Rotary heat exchanger



y 4, 1966 G. J. HUEBNER, JR 3,252,505

ROTARY HEAT EXCHANGER Filed June 22, 1962 5 Sheets-Sheet 1 INVENTUR. Georye flzze'zrzef, Jr!

irranvvz' si 1965 G. J. HUEBNER, JR 3,252,505

ROTARY HEAT EXGHANGEH Filed June 22, 1962 5 Sheets-Sheet 2 INVEN TOR.Geo rye J' b zzeirzer; J7?

BY ALA/1m wh m May 24, 1966 G. .1. HUEBNER, JR

ROTARY HEAT EXCHANGER 5 Sheets-Sheet 3 Filed June 22, 1962 "4 4 a 1 a. Vz 7 2 d g Q w 2 y M 4 z 4 7 4 x IN VEN TOR. fieo rye flael/ver; BY

May 24, 1966 G- J. HUEBNER, JR

ROTARY HEAT EXGHANGER 5 Sheets-Sheet 5 Filed June 22, 1962 Nudneiz (Vz/rref: Far 717M227; 407117740" 770w United States Patent This applicationis a continuation-in-part of copending application Serial No. 379,673,filed September 11, 1953, now abandoned.

This invention is concerned with improvements in a I regenerator of thecounterflow type for an automotive gas turbine engine wherein thedifficult problems of reducingweight and size are of paramountimportance and must be considered concurrently with the problems ofminimizing resistance to gas flow through the regenerator matrix whileat the same time achieving rapid and efiicient heat transfer from thehot exhaust gases to the cooler high pressure inlet air from thecompressor.

According to one embodiment of the invention as set forth in thefollowing specification, a rotary heat exchanger having a generallycylindrical form is provided with a hub upon which is formed a glassmatrix struc ture having smooth straight axially extending passagestherethrough. A peripheral rim member is mounted upon the glassstructure in concentric relationship with respect to the hub. Theassembly is mounted transversely with respect to the air discharged froman intake compressor unit and with respect to the exhaust passageways ofa gas turbine powerplant. The axial passages through the glass matrixstructure are adapted to'conduct the compressed intake air and exhaustgases there/through at angularly spaced positions. The exhaust gases areeffective to heat the proximate glass structure to a temperature whichapproaches in magnitude the temperature of the powerplant exhaust gases.Upon rotation of the regenerator about its axis, the heated portion ofthe regenerator glass matrix is brought into contact with the coolercompressed intake air as the same passes through the heated airpassages, thereby heating the compressed gas turbine intake air which isthen conducted to the burner.

One of the several objects of the present invention is to provide arotary regenerator unit which will be capable of establishing atemperature gradient across the length of the gas passages therein. Thetemperature gradient is highly desirable because it makes possible agreater increase in the temperature of the intake air.

Another object of the present invention is to provide a rotaryregenerator unit for use with a gas turbine powerplant or the like whichis resistant to corrosion during high temperature operating conditons.

Another object of the present invention is to provide a rotaryregenerator unit for use with a gas turbine powerplant or the like whichis adapted to retain uniform dimensions and to resist the tendency towarp while it is being heated by the turbine exhaust gases. Thisantiwarping feature is of considerable importance in that the intake airmust be effectively sealed from the exhaust gases in order' to maintainoperating efliciency at a relatively high value. Working of theregenerator greatly impairs the effectiveness of the sealing structure.

Another object of the present invention is to provide a rotaryregenerator unit for use with a gas turbine powerplant or the like whichis relatively light in weight and inexpensive to manufacture and whichachieves optimum heat transfer efficiency with minimum size and weight.

Although the counterflow type regenerator itself is not new, such as aregenerator wherein comparatively cool inlet air flows in one directionthrough parallel gas passages of the matrix, followed by flow of the hotexhaust gases in the opposite direction through the same passages, givesrise to particular problems in the attainment of efiicient operationwhich are not encountered in other types of heat exchangers. Animportant aspect of the present invention has been the discovery that aparticular relationship between the thickness of the walls of theindividual gas passages and the shape and hydraulic diameter of thesepassages results in a regenerator of optimum heat transfer efiiciencywith minimum size and weight. In particular, it has been discovered thatthe reduction in thickness of the individual gas passage walls goeshand-in-hand with an appreciable elongated cross sectional shape inorder to obtain optimum heat transfer.

'In order to concisely set forth the structural characteristics and modeof operation of the present invention, reference will be made to theaccompanying drawings in which:

FIGURE 1 is a schematic representation of an automobile gas turbinepowerplant showing the relative position of the various componentsthereof including the rotary regenerator of the present invention;

FIGURE 2 is an end view of the assembly of FIG- URE 3;

FIGURE 3 is a complete assembly view of an actual operative gas turbinepowerplant showing the rotary heat exchanger of the present inventionincorporated therein;

FIGURE 4 is a View taken along section line 4-4 of FIGURE 3 showing thesegmental inlet and outlet portions for conducting air through theregenerator in either axial direction;

FIGURE 5 is a perspective view partly in section showing the rotaryregenerator unit of the present invention isolated from the assembly ofFIGURE 3;

FIGURE 6 is a detail enlarged view showing the structure of the glassmatrix of FIGURE 5;

FIGURE 7 is an enlarged schematic cross sectional view showing the heatflow associated with a single gas fiow passage;

FIGURE 8 illustrates the relationship between Nusselts number N and afunction =K T/K D for various regenerator flow passages of the typeillustrated in FIGURE 7; and

FIGURE 9 illustrates the relationship between Nusselts number and thewall thickness of various ceramic regenerator flow passages of the typeillustrated in FIG- URE 7.

For a more complete description, reference will be made first to theschematic representation of a gas turbine powerplant in FIGURE 1 whereinnumeral 10 is used to designate the gas turbine compressor with an inletat 11 and a'high pressure discharge passage at 12. The passage 12conducts intake air to the rotary regenerator 13 which in turn allowsthe intake air to pass axially therethrough into passage 14 and theburner 15. Upon reaching the burner, the intake air is mixed with fueland then burned. The burned gases pass through passage 17 to the turbine18 which in turn drives the compressor shaft 19 and the power outputshaft 20. The hot turbine exhaust gases are then conducted throughpassage 21 and the regenerator 13 to the exhaust port 22. The exhaustgases cause a portion of the regenerator to become heated while passingtherethrough. A suitable driving means 23 is provided to rotate theregenerator about its own axis which causes the portion of theregenerator unit which is heated by the exhaust gases to come in contactwith the intake air passing from passages 12 to 14 thus heating the sameto a temperature which approaches in value the temperature of theexhaust gases.

Having thus described the general arrangement of the component elementsof the powerplant, reference will now be made to the operativepowerplant assembly as shown in FIGURE 3 which incorporates theregenerator unit of the present invention. The compressor unit, showngenerally at 10, comprises a converging inlet 11 which is open to thesurrounding atmosphere throughout its entire periphery. A rotor member24 of the compressor is effective to force the intake air from the inletportion 11 into an annular diffuser member 29 where a suitable intakeair pressure head is created. The rotor member 24 comprises suitablevanes 25 disposed about the hub 28 of the rotor for forcibly feeding theintake air to the diffuser 29 and also a radially extending portion 26having a peripheral discharge opening 30 for creating a centrifugalhead. The diffuser 29 blends into a wide mouthed port-ion, shown at 31,as it progresses circumferentially about the axis of the powerplant. Theportion 31 covers a segmental portion of the regenerator unit shown at13. This segmental portion is shown at X in FIGURE 4. A graphite sectorplate or seal 32 is provided for sealing the inlet portion 31. Smoothcontact surfaces 32' and 32 are provided on seal 32 for engaging therotating surface of the regenerator unit 13 and the diffuser portion 31respectively. An additional sealing means may be provided at 33 toinsure a sealing contact between portions 31 and the stationary graphiteseal 32. Another graphite sector plate or seal 35 is provided on theopposite side of the regenerator unit and is shaped similar to seal 32.Both of the seals 32 ad 35 have segmental openings to allow the portionX of the regenerator unit to be exposed.

The passage 14, which was referred to previously in connection with theschematic drawing of FIGURE 1, is shown in FIGURE 3 in close proximityto the regenerator unit 13 and is adapted to conduct the intake airpassing through the regenerator from diffuser portion 31 to the burner15. After the fuel is mixed with the intake air and burned, the burnedgases pass into the passage 36 and then into the exhaust chamber 42.While passing from passage 36 to chamber 42 the burned gases passthrough an annular channel 37 in which are disposed annular cascades ofturbine blades 38 and 39 and associated stator blades 40 and 41. Theburner 15 is illustrated by means of an outer elevation view in FIGURE2. The actual path followed by the intake air and combustion gasesthrough the burner will not be described in detail since the burner doesnot form a part of this invention.

The blades 38 are mounted upon a primary turbine member 43 which ismounted upon an axially extending shaft assembly 44 which in turn isdrivably secured to the compressor rotor hub 28. The blades 39 aremounted upon a secondary turbine member 45 which is mounted upon anaxially extending shaft 46. The shaft 46 is drivably connected to theinput pinion gear of a reduction gear transmission assembly which isdesignated generally by numeral 47. A power absorbing means may beconnected to the transmission power output tails'haft 48.

The exhaust chamber 42 is formed to cover a large segmental portion ofthe regenerator unit 13 which is shown at Y in FIGURE 4. The graphiteseals 32 and 35 are also provided with mating segmental openings tocause the portion Y of the regenerator to be exposed to the exhaustgases. A suitable housing 50 is adapted to form a passageway 51 toconduct the exhaust gases passing through the segmental portion Y of theregenerator to an exhaust port.

It should be noted from FIGURE 3 that the seals 32 and 35 are effectiveto seal the exhaust gases passing from chamber 42 to passageway 51 aswell as to seal the intake air passing from portion 31 to passageway 14so that the exhaust gas and intake air are not intermixed.

The total area provided at Y for the passage of the exhaust gasesthrough the regenerator unit is larger than the area at X for the intakeair because the volume occupied by the heated exhaust gas is necessarilylarger. It

has been found that a ratio of approximately two to one between theexhaust and intake areas is adequate for the usual operatingtemperatures encountered.

The regenerator unit 13, which comprises the subject matter of thepresent invention, is shown more in detail in FIGURES 5 and 6. The unititself comprises a hub 52 and a rim portion 53 which may be made of anysuitable material, such as steel. The hub may be provided with asuitable bearing means, as shown in FIGURE 3 at 54, for rotatablymounting the regenerator unit upon a shaft 55 which in turn is suitablymounted in the outer housing. The rim portion is provided with aperipheral ring gear 56 which serves as a means for driving theregenerator unit about its own axis on bearing means 54. Any suitablepower source may be used to drivably engage the ring gear 56.

The body of the regenerator unit 13 comprises a glass matrix structurewhich is formed upon the hub 52 and secured thereon by means of theconcentric rim portion 53. The term glass herein includes any smoothsurfaced glass-like or ceramic material capable of withstanding thecyclic temperature extremes to which the regenerator is subject duringoperation and having a high coefficient of specific heat and lowcoefficients of thermal conductivity and expansion. The glass matrixstructure, as seen in FIGURES 5 and 6, comprises a thin glass strip orribbon 60 having a thickness of approximately .004 inch, but which mayvary as described below in accordance with FIGURE 8, depending on thethermal and strength properties of the specific material. A series oftransverse ribs 61 are formed on the glass strip at spaced intervals ofapproximately .150 inch. It is desirable to form the ribs with a heightof approximately .012 inch. The strip 60 is wound about the hub 52 sothat the ribs are effective to maintain a clearance between the layersas shown in FIGURE 6. The layers of glass strips may be fused togetherduring the winding operation or they may be fused together as a unitafter the winding has been completed.

The axial thickness of the glass matrix may be varied according to thedesign requirements. It has been found, however, that a thickness ofapproximately three inches is adequate when the unit is used with a gasturbine power plant having a rated power of about horsepower. The otherdimensions may also be varied as described in more detail below in orderto adapt such a regenerator unit to a particular application.

For the purpose of particularly pointing out the mode of operation andthe effectiveness of the present invention, the path of gas flow throughthe gas turbine power plant will be followed together with a referenceto some typical operating temperatures for a 150 horsepower power plant.An attempt has been made to illustrate the path of gas flow by means .ofarrows in FIGURE 3.

It will be assumed that standard atmospheric conditions exist at theintake portion of the compressor 10. When the intake air reaches therotor 24 of the compressor, the temperature is still about the same asthe temperature of the surrounding atmosphere. When the air passesthrough the rotor and is compressed, the temperature increases toapproximately 400 F. The temperature of the intake air is thereforeapproximately 400 F. at the time it enters the inlet segmental portion Xof the regenerator unit 13. While passing through the regenerator fromchamber 31 to passage 14, the intake air temperature is increased toabout 900 F. The passageway 14 conducts the heated intake air to theburner 15 where the combustion process causes the temperature toincrease to about 1500 F. The products of combustion are conductedthrough passage 36 to the turbine members. The work performed on theturbine member is accompanied by a temperature drop to approximately1000 F. The heated exhaust gases then pass from chamber 42 through thesegmental portion Y of the heat exchanger unit and heat up the glassmatrix. This is accompanied by a temperature drop in the exhaust gasesto approximately 500 F. The passageway 51 then conducts the cooledexhaust gas out a suitable exhaust port.

It should be observed that the hot exhaust gases pass from one side ofthe heat exchanger to the other in an axial direction. Accordingly,because of the low thermal conductivity of the matrix, a temperaturegradient will become established across the axial thickness of the unitwith the higher temperature existing at the gas inlet side which isclosest to chamber 42 and the lower temperature existing at the gasoutlet side. The regenerator unit is constantly rotated during theoperation of the power plant. Accordingly, the hotter part of the matrixwill be in contact'with the intake air outlet side when rotated so as tointersect the intake air stream. The cooler part of the matrix will, ofcourse, come into contact with the intake air inlet side. Because of thetemperature gradient, it is possible to heat the intake air to a highervalue than that which would result if such a gradient did not exist,since the intake air comes in contact with portions of the matrix whichare heated to a temperature considerably greater than the meantemperature. Since the overall thermal efficiency of the engine isdirectly dependent upon the temperature of the intake air, it ispossible to obtain a higher overall efiiciency for the entire unit bymaking use of the regenerator unit of the present invention.

Prior to the present invention, little was known to the art regardingthe heat transfer properties of a multitude of small parallel gaspassages having thin walls of the character and dimensions disclosedherein. In order to facilitate understanding of some of the problemsinvolved, an enlarged cross sectional view of a single gas passage 62 isillustrated in FIGURE 7 wherein two of the heat flow components areindicated by arrows A and B.

Hot exhaust gas flowing through passage 62 in a direction perpendicularto the plane of the paper will transfer heat to the inner side walls asindicated by the arrows A. It has long been recognized that the thinnerthe walls 60 and 61, the greater will be the surface area for any givenweight of material and, as far as this factor is concerned, the greaterwill be the total heat transfer from the gas to the passage walls in agiven time limit.

In order to take advantage of the thin-walled effect, the obvious stepwas to employ a fibrous matrix, as for example glass or metal fibers forthe heat exchange medium. Such constructions have been unsuitable foruse with automotive gas turbine engines wherein the gas flow is atcomparatively high velocity and pressure. Not only do particles of thefibrous material break off when subject to the high pressure and cyclictemperature changes and damage the extremely high speed turbine blades,but the resistance to gas flow is prohibitive in a regenerator havingrandom gas flow passages.

Accordingly, the present invention utilizes a multitude of preformedflow passages having thin smooth walls extending directly through theregenerator in axial side-byside relationship and sealed along theiraxial length to prevent circumferential flow of gases from one gaspassage to another within the regenerator matrix. The use of the smoothaxial flow passages enables an increase in the total surface area to anoptimum value for any given weight of regenerator without undulyincreasing the flow resistance. It will also be assumed herein that theregenerator matrix is designed for laminar allow of the gasestherethrough, as distinguished from turbulent flow, for the sake ofminimizing the resistance to the gas flow. In addition as described morefully below, the dimensions specified achieve optimum thermal efficiencyby providing answers to problems that were not known to existheretofore.

Referring to FIGURE 7, it has been found that the transfer of heat fromthe gas to the passage walls of elongated cross sectional area issubstantially less adjacent the small edges a than at the mid-region ofthe long dimension [2 of the passage. Thus each flow passage 62 adjacentits small edges a will be relatively cooler than said mid-regions. Inconsequence, the portions of each gas passage 62 adjacent its oppositesmall edges a will be relatively ineffective as a heat transfer mediumunless the heat flow indicated by the arrows B within the material ofthe walls can be used to conduct heat within the passsage side wallstoward -the small edges a.

It is therefore to be realized that a definite but heretofore unexpectedrelationship exists between cross sectional elongation and Wallthickness. The less elongation and the greater number of total gaspassages, the greater will be the total effect of the small edges a inreducing the heat transfer efficiency of the regenerator. Where thewalls are reduced to a film-like thickness, as in the present invention,the cross sectional elongation must be increased substantially,otherwise the loss in heat transfer efiiciency resulting from reducedheat transfer toward the small edges of each passage becomes asignificant consideration.

On the other hand, the thinner the passage walls, the greater will bethe resistance to heat flow therein toward the small edges a and thegreater will be the loss in heat transfer efficiency. This latterconcept is directly opposed to the practice heretofore of attempting toincrease the total surface area and heat transfer efficiency bydecreasing the wall thickness of the individual gas passages. As

explained below, the gas passage walls cannot be reduced in thicknessbelow a predetermined minimum without sacrificing optimum heat transferefficiency.

As the heat flow in the direction of the arrows B increases, FIGURE 7,the effective utilization of the small end regions of the elongated gasflow passages and correspondingly the heat transfer efiiciency of theregenerator will be increased. Likewise, as the elongation of the crosssectional area of each gas passage is increased, the total effect of thesmall edge portions a of all the passages tending to reduce theregenerator heat transfer efficiency will be decreased.

If we assume side walls 60 of infinite thickness so as to minimizeresistance to heat flow in consequence of the wall thickness factor, amodulus of heat transfer effectiveness well known as Nusselts number canbe correlated directly with the heat flow within the walls of the gaspassages in the directions of the arrows B. For a gas flow passage ofsquare cross section, Nusselts number equals 3.6.

For an elongated gas passage having side Walls in the ratio 'to 12 asspecified by applicant, Nusselts number increases to slightly less than7. Thereafter as the elongation increases to infinity (parallel sidewallswith no small end walls) Nusselts number increases to approximately8.2. The ratio of the dimensions b/a=G is hereinafter referred to as theaspect ratio. Thus as the elongation or aspect ratio G increases from asquare or 1:1 ratio to 150:12 ratio, Nusselts number rapidly doubles invalue. As the elongation is increased infinitely, Nusselts numberasymptotically approaches 8.2. The ratio 150112 is accordingly within acritical range in that it is at the region of optimum Nusselts numberfor any practical obtainable increase in the elongation or aspect ratioG. Below the optimum aspect ratio, Nusselts number decreases rapidly.Above the optimum aspect ratio, any advantage obtained from an increasein Nusselts number is negligible whereas the feasibility of increasingthe aspect ratio rapidly decreases because of the lack of rigidity ofthe thin walls which cannot be maintained in parallel spacedrelationship over extended distance without intermediate support.

Refering to FIGURE 8, Nusselts number N =HD/K is plotted against adimensionless parameter =K T/K D. In the above expressions:

H is the coefficient of thermal conductance between the walls of the gaspassage 62 and the gas flowing therein and measures the quantity of heatflow between the gas and a unit area of the sidewalls per unit time andtemperature diflerential.

D is the hydraulic diameter of the'elongated passage 62 and equals fourtimes the flow area divided by the perimeter, which for a rectangle is2ab/(a-t-b).

T is the thickness of the sidewalls b, the thickness of the walls abeing unimportant when the aspect ratio G=b/a is large. Both T and D arelinear measurements, expressed for example in feet (ft).

K and K are the coeflicients of thermal conductivity of the sidewallsand gaseous fluid medium respectively and measure the quantity of heatflow per unit time and temperature differential along a unit distancewithin the sidewalls and gaseous medium respectively. Inasmuch as thecombusition products of an automotive gas turbine engine comprise acomparatively small portion of the total mass of the inlet air, thedifference in the value of K: for the inlet air and exhaust gas isnegligible. Hence the value of K for air is feasibly employed herein.

It is apparent that Nusselts number N is also dimensionless because,employing conventional units, H /K: can be expressed as is thereforedimensionless. The family of curves in FIG- URE 8 accordingly representthe relationships between two dimensionless parmeters at various aspectratios and are entirely independent of the material employed for thewalls of the gas passage 62 and the fluid flowing therethrough.Regardless of the composition of the fluid medium or the material of theregenerator matrix, there is a definite value of N for each value of andG.

For a regenerator of any given size, the total heat transfer area of thematrix will be increased by increasing the number of gas passages 62.Thus the hydraulic diameter D of each passage 62 will be as small aspracticable in order to obtain the maximum number of gas passages 62,the smallness of the hydraulic diameter being limited by the total axialpressure drop across the regenerator that can be tolerated and thevolume of gas that can be accommodated by laminar flow. Inasmuch as as Nincreases in value, the conductance H and correspondingly the efficiencyof the regenerator increases. Similarly, g5 and correspondingly theweight of the regenerator will increase in value with increasing wallthickness T by reason of the expression =K T/K D.

It is apparent from FIGURE 8 that Nusselts number N increases rapidly asthe aspect ratio G increases to the practicable limits 10 or 12, asexplained above, but decreases rapidly with decreasing 5 at valuessmaller than approximately 10, which latter value defines the knee orapproximate mid-point of the sharp bend of each curve. For larger valuesof gb, the thickness T of the passage sidewalls can be increasedinfinitely without appreciably enhancing the value of Nusselts number N.

On the other hand, as indicated by the two vertical 10% measurementsassociated with 6:10 and G=12 respectively in FIGURE 8, 90% of thetheoretical maximum value of Nusselts number N is obtained on the 6:10curve if 5 is approximately 4, and is obtained on the G=12 curve if isapproximately 5. The optimum value of q for minimum regenerator mass andmaximum heat transfer effectiveness as measured by Nusselts number N istherefore approximately 10, (log =1) but values of 5 as low as 4 (log=.6) will achieve more than 90% of the value of Nusselts number if theaspect ratio G is greater than 10; and similar values of N are obtainedwhen log p is 1.4 and G is 6. The two vertical 10% measurements indicatethe range of within the upper 10% of the maximum theoretical values of Nassociated with the curves G=10 and 6:12 respectively. N decreases veryrapidly as G decreases toward 4. The lower practical limit for G from athermo-dynamic standpoint is thus somewhere between 4 and 6 and may bestated as being on the order of about 5.

Specific values of gas passage wall thickness for glass or ceramic whereG=10 and Kf/Kw'- .026 are illustrated for various hydraulic diameters inFIGURE 9 to emphasize the sharp drop in heat transfer effectiveness whenthe wall thickness is decreased below the critical value. The uppermostcurve D=.025" corresponds approximately to the hydraulic diameter of thegas passage 62 in FIG- URE 7, wherein the wall thickness of the ceramicgas passages ranges between approximately .0025" and .006 for values of5 between approximately 4 and 10. The optimum wall thickness for othersuitable matrix materials having the desired high specific heat and lowcoefficients of thermal conductivity and expansion will similarly bedetermined by the value of in the critical range log 6:11.4. The maximumhydraulic diameter D equal to .06" may be calculated by substituting theminimum value of equal to 4, the maximum wall thickness T equal to.006", and the value of [(j/Kw equal to .026 in the expression: equals KT/K D.

I claim:

1. A counterflow type regenerator matrix for an automotive gas turbineengine wherein high temperature exhaust gases and cooler inlet gases arealternately directed through portions of said matrix in oppositedirections by substantially laminar flow, said matrix comprising acomparatively rigid ceramic body having a multitude of substantiallyparallel gas passages of elongated cross sectional area, the walls oflonger cross section of each gas passage being substantially parallel toeach other and sufliciently smooth and impervious to said gases toeffect said laminar flow and being at least several times longer thanthe smaller cross sectional dimension of said passage, the hydraulicdiameter of each gas passage being not greater than a value on the orderof about .06", and log K T/ KfD being between values on the order ofabout .6 and 1.4, wherein K, is the coefficient of thermal conductivityof. the gas flowing through said gas passages, K is the coeflicient ofthermal conductivity and T is the thickness of said walls of longercross section, and D is the hydraulic diameter of said passages.

2. The generator matrix as in claim 1 wherein the spacing between saidwalls of longer cross section is on the order of about .01".

3. The regenerator matrix as in claim 2, wherein the ratio of the longdimension to the short dimension of the cross section of each gaspassage is not less than a value on the order of about 5.

4. The regenerator matrix as in claim 1, wherein the ratio of the longdimension to the short dimension of the cross section of each gaspassage is not less than a value ranging on the order of from about 10to 6 while log K T/K D ranges correspondingly on the order of from about.6 to 1.4.

5. The regenerator matrix as in claim 4 wherein the spacing between saidwalls of longer cross section is on the order of about .01".

6. A counterflow type regenerator matrix for an automotive gas turbineengine wherein high temperature exhaust gases and relatively coolerinlet gases are alternately directed through portions of said matrix inopposite directions by substantially laminar flow, said matrixcomprising a comparatively rigid body of ceramic material having acomparatively high coefiicient of specific heat and a comparatively lowcoefiicient of thermal conductivity and being formed with a multitude ofsubstantially parallel gas passages of elongated cross sectional area,the walls of longer cross section of each gas passage beingsubstantially parallel to each other and suflieiently smooth andimpervious to said gases to effect said laminar fiow and being at leastseveral times longer than the smaller cross sectional dimension of saidpassage, and log KWT/KID being on the order of about 1, the hydraulicdiameter of each gas passage being not greater than a value on the orderof about .06, wherein K is the coefficient of thermal conductivity ofthe gas flowing through said gas passages, K is the coeflicient ofthermal conductivity and T is the thickness of said walls of longercross section, and D is the hydraulic diameter of said passages.

7. The counterflow type regenerator matrix as in claim 6, saidcoefficients of specific heat and thermal conductivity being comparableto the corresponding coefiicients of high temperature resistant glass.

8. The counterflow type regenerator matrix as in claim 7, the spacingbetween the walls of longer cross sectional dimension being on the orderof about .01.

9. The counterflow type regenerator matrix as in claim 8, the ratio ofthe longer cross sectional dimension to the shorter cross sectionaldimension of each gas passage being not less than a value on the orderof about 5.

10. The counterflow type regenerator matrix as in claim 6, the hydraulicdiameter of each gas 'passage being on the order of about .02".

11. The counterflow type regenerator matrix as in claim 10, the ratio ofthe longer cross sectional dimensiOn to the shorter cross sectionaldimension of each gas passage being not less than a value on the orderof about 5.

12. In combination, a counterflow type regenerator of an automotive gasturbine engine for transferring thermal energy from a stream of heatedexhaust gases to a generally oppositely directed stream of relativelycooler inlet gases, said regenerator comprising a comparatively rigidceramic matrix having a multitude of substantially parallel gas passagesof elongated cross section adapted for substantially laminar flow ofsaid gases therethrough, the Walls of longer cross section for each gaspas-sage being substantially parallel to each other, and sufficientlysmooth and impervious to said gases to effect said laminar flow, theratio of the long dimension to the short dimension of the cross sectionof each gas passage being not less than a value on the order of about 5,the hydraulic diameter of each, gas passage being not greater than avalue on the order of about .06, means for conducting the two oppositelydirected streams of gases alternately to a portion of said matrix forsaid laminar flow thereth-rough, and log K T/K D being between values onthe order of about .6 and 1.4 wherein K and K are the coefiicients ofthermal conductivity of said walls of longer cross section and the gasflowing through said passages respectively, T is the thickness of saidwalls of longer cross section, and D is the hydraulic diameter of eachgas passage.

References Cited by the Examiner UNITED STATES PATENTS 2,023,965 12/1935Lysholm 165-1 2,552,937 5/1951 Cohen 1651O 2,596,642 5/1952 Boestad165-166 2,706,109 4/1955 Od-rnan 165l0 ROBERT A. OLEARY, PrimaryExaminer. CHARLES SUKALO, Examiner.

R. E. BACKUS, Assistant Examiner,

1. A COUNTERFLOW TYPE REGENERATOR MATRIX FOR AN AUTOMOTIVE GAS TURBINEENGINE WHEREIN HIGH TEMPERATURE EXHAUST GASES AND COOLER INLET GASES AREALTERNATELY DIRECTED THROUGH PORTIONS OF SAID MATRIX IN OPPOSITEDIRECTIONS BY SUBSTANTIALLY LAMINAR FLOW, SAID MATRIX COMPRISING ACOMPARATIVELY RIGID CERAMIC BODY HAVING A MULTITUDE OF SUBSTANTIALLYPARALLEL GAS PASSAGES OF ELONGATED CROSS SECTIONAL AREA, THE WALLS OFLONGER CROSS SECTION OF EACH GAS PASSAGE BEING SUBSTANTIALLY PARALLEL TOEACH OTHER AND SUFFICIENTLY SMOOTH AND IMPERVIOUS TO SAID GASES TO EFECTSAID LAMINAR FLOW AND BEING AT LEAST SEVERAL TIMES LONGER THAN THESIMILAR CROSS SECTIONAL DIMENSION OF SAID PASSAGE, THE HYDRAULICDIAMETER OF EACH GAS PASSAGE BEING NOT GREATER THAN A VALUE ON THE ORDEROF ABOUT .06", AND LOG10 KWT/ KFD BEING BETWEEN VALUES ON THE ORDER OFABOUT .6 AND 1.4, WHEREIN KF IS THE COEFFICIENT OF THERMAL CONDUCTIVITYOF THE GAS FLOWING THROUGH SAID GAS PASSAGES, KW IS THE COEFFICENT OFTHERMAL CONDUCTIVITY AND T IS THE THICKNESS OF SAID WALLS OF LONGERCROSS SECTION, AND D IS THE HYDRAULIC DIAMETER OF SAID PASSAGES.